Planetary transmission mechanism



June 14, 1938. w COTTERMAN 2,120,832

PLANETARY TRANSMI S S ION MECHANI SM Original Filed Sept 17, 1935 4 Sheets-Sheet 1 I //VVE/V70/ 249M Wa'm June 1938. F. w. COTTERMAN PLANETARY TRANSMISSION MECHANISM Original Filed Sept. 17, 1935 4 Sheets-Sheet 2 lai June 14, 1938. F w COTTERMAN 2,120,832

PLANETARY TRANSMISS ION MECHANISM Original Filed Sept. 17, 1935 4 Sheets-Sheet 5 III III June 14, 1938. F. w. COTTERMAN 2,120,832

PLANETARY TRANSMISSION MECHANISM Original Filed Sept. 17, 1935 4 Sheets-Sheet 4 Poll/10.5 $77555 [PM 0; 14 516,973

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Patented June 14, 1938 UNITED STATES PLANETARY TRAN SMIS SION MECHANISM Frederick W. Cotterman, Dayton, Ohio, assignor of one-half to Bessie D. Apple, Dayton, Ohio Original application September 17, 1935, Serial No, 40,946; Divided and this application August 3, 1936, Serial No, 94,053

20 Claimso This invention is a division of my copendlng application Serial No. 40,946, filed September 1'7, 1935, and relates to a power transmission mechanism, particularly applicable to automotive vehicles and embodies some of the features of my copending application Serial No. 26,765, filed June 15, 1935.

An object of the invention is to provide transmission mechanism in which the most used or middle portion of the driving range of a vehicle will be efiected through direct drive, to the end that the only gearing normally operative between the engine and road wheels will be the conventional axle gearing, which is usually either a worm and wheel, or a bevel pinion and gear, then supplementing the axle gearing with a speed reducing or underdrive gear-set between the engine and axle automatically operative at the lower speeds upon overload, and with a speed increas ing or overdrive gear-set between the axle gearing and the road wheels which may become automatically operative only at exceptionally high speeds.

Another object is to provide a fluid coupling 25 for connecting the underdrive gear-set to the engine, whereby less speed reduction need be had through the underdrive gear-set and consequently less engine racing results in accelerating the vehicle to a given speed.

Another object is to provide, for the underdrive gear set, one having constant mesh gearing but wherein there are no roller, spring, or friction clutches, but only two jaw clutches, one, the gear drive clutch, being operative when the power is being transmitted through the gearing and the other, the direct drive clutch, when the power is being transmitted through the mechanism directly, the jaw clutches being provided with means whereby they may become engaged and disengaged only at synchronism of their respective parts to the end that they may engage and. disengage at high vehicle speeds without rattle or clash.

Another object is to provide, in the underdrive gear-set, clutch engaging means which at all times tends to hold the direct drive clutch in engagement, with torque responsive means tending to disengage said clutch upon overload, and with speed responsive means for increasing the capacity of the clutch engaging means to hold the clutch engaged which is more nearly in proportion as the square root of the R. P. M. instead of as the square of the R. P. M. as in common practice, to the end that the overload point of the clutch may be sufilciently high at the lower speedsv without being too high at the higher speeds.

Another object is to provide, in the underdrive gear-set, torque responsive means which will, upon disengagement of the direct drive clutch by overload, hold the said clutch completely and fully disengaged until the torque being applied is substantially reduced by decrease of the fuel being suppliedto the engine.

Another object is to provide, in the overdrive gear-set which is interposed between the conventional axle gearing and the road wheels, two speed responsive jaw clutches, one of which normally connects the said axle gearing and road wheels directly independently of the overdrive gearing, the other being adapted to become operative at a predetermined speed, by momentary release of the applied torque, to connect the said axle gearing to the road wheels through the overdrive gearing, to the end that the lesser percentage of driving only which is done at very high speeds need be done through the overdrive gearing thereby leaving all normal speeds to be effected through the axle gearing only.

Another object is to so construct the overdrive speed responsive jaw clutches as to insure that, in the transition from direct to overdrive connection or vice versa, there will be no interval wherein neither clutch is engaged, to the end that no overrunning clutches, such as roller or spring clutches need be used and no free wheeling will be had either in direct, in overdrive, or in the transition period between the two drives.

Another object is to provide means whereby the same overdrive gear-set may be made operative in a reverse direction as speed reducing gearing for reversing the vehicle, to the end that no additional gears need be provided for this purpose.

Another object is to provide a manually operable means operative to three positions to provide forward, neutral and reverse connections between the axle gearing and the road wheels, said means being incorporated in the axle mechanism and contained within the axle housing.

That these and many other objects and meritorious features are attained will become apparent as the mechanism is described in greater detail and reference is had to the drawings, wherein;

the mechanism taken on the lines l--l of Figs. 2 and 3.

Fig. 2 is a longitudinal horizontal section'taken on the line 2-2 of Fig. 1.

,. Fig. 1 is a longitudinal vertical section through 0 Fig. 3 is a horizontal section taken on the line 3-3-of Fig. 1.

Fig. 4 is a transverse section taken at 4-4 of Fig. 3 through the speed responsive device which operates the overdrive clutches.

Fig. 5 is a transverse section taken at 5-5 of Fig. 3 through the overdrive clutch parts.

Fig. 6 is a transverse section taken at 8-5 of Fig. 3 through the overdrive gearing.

Fig. 7 is a transverse section taken at 'I-| of Fig. 3 through the manually operable forward, rearward, and neutral mechanism.

Fig. 8 is a flattened section along the arcuate surface 88 of Figs. 3 and 4 showing the stationary and movable parts of the speed responsive dcvice which operates the overdrive clutches.

Fig. 9 shows diagrammatically a series of the clutch teeth of Fig. 5 to illustrate the manner in which the side faces of the clutch teeth must be beveled.

Fig. 10 is a section taken at Ill-l of Fig. 2 through the operating rod and lever of the manually operable forward, rearward, and neutral mechanism showing the detent mechanism for holding the device in its several positions.

Fig. 11 shows diagrammatically a series of the I teeth of the direct drive clutch of the underdrive gear-set showing how parts of the teeth of the driving member are kept misaligned for the purpose of excluding the teeth of the driven member until synchronization of the driving and driven members takes place.

Fig. 12 is a view similar to Fig. 11 except that the clutch teeth are shown with the parts of the teeth of the driving member aligned to receive the teeth of the driven member.

Fig. 13 is a transverse section taken at l3-l3 of Fig. 1 through the driving element of the underdrive gear-set illustrating .also the driving lugs of the driving member of the direct drive clutch.

Fig. 14 is a transverse section taken at l4|4 of Fig. 1 through the underdrive gearing and through the springs which constitute the variable means for keeping the direct drive clutch engaged, showing also the parts of the clutch driving teeth in the misaligned non-engageable arrangement.

Fig. 15 is a transverse section taken at |l5 of Fig. 1 showing the greater portion of the underdrive gear-set including its speed responsive operating device in end elevation.

Fig. 16 is a detail perspective view of the sun gear of the underdrive gear-set showing its integral jaw clutch part which, upon engagement holds the sun gear stationary thereby to effectuate gear drive.

' Fig. 17 is a detail perspective view of the stationary jaw clutch part which engages the part in Fig. 16 to hold the sun gear against rotation.

Fig. 18 is a fragmentary section taken at I8l 8 of Fig. 15 showing the several parts of the direct drive clutch, its pressure pins, pressure plate, cinch engaging springs, and spring stressing p e.

Fig. 19 is a fragmentary section taken at l9l 9 of Fig. 15 showing one of the guide pins of the spring pressure plate and a centrifugal weight and cam for moving the spring pressure plate.

Fig. 20 is a fragmentary section taken at 28--20 of Fig. 15 showing the means of securing together the two parts of the planet pinion carrier which is the driven element of the underdrive gear-set.

Fig. 21 is a diagram showing a number of positions which the center of gravity of one of the centrifugal weights takes and corresponding positions of the cam which regulates the stress of the direct drive clutch engaging springs.

Fig. 22 is a diagram showing the clutch engaging force of the springs at various vehicle speeds.

The crank shaft 38 of an internal combustion engine carries a fluid coupling comprising the flywheel 32 to the outer face of which the cover 34 is secured by screws 36. The cover 34 carries the driving vanes 38 and a hub 40 having a bearing bushing 42 within which the driven element rotates.

The driven element 44 carries the vanes 48 and the central hollow journal 48 upon which the driven member has rotative hearing. The journal 48 is internally splined to receive the externally splined drive shaft50 of the underdrive gear-set. A ball bearing 52 is provided to take the thrust of the driven element 44.

A flywheel cover 54 encloses the flywheel and coupling and supports the transmission housing. The outer face of the cover 54 is closed by the end wall 56 which serves also as an end closure for the transmission housing. A main casting 58 is screwed to the wall 56 by the screws 60. Casting 58 integrally comprises the transmission housing 62 and the axle housing 64, the housing 62 containing the underdrive gear-set and the housing 64 containing the axle gearing, which in this case consists of a worm and a worm wheel, and the overdrive gear-set.

An end head 66 closes one end of the axle housing 64 and is secured thereto by the screws 85 and provides space for the overdrive clutches and.

their speed responsive clutch operating mechanism. A bearing plate 68 separates the space within the axle housing 64 from the space within the end head 66.

A second end head closes the other end of the axle housing 64 and is secured thereto by the screws 65 and provides space for the manually operable forward, rearward and neutral mechanism.

Both the underdrive and overdrive gear-sets herein employed are of the planetary type each comprising a sun gear, several planet pinions adapted to both rotate upon their axes and revolve around the sun gear, a planet pinion carrier comprising means upon which the pinions may rotate about their individual axes and means compelling revolution of the pinions bodily in an orbit concentric with the sun gear, and a ring gear surrounding and meshing with the planet pinions.

In the underdrive gear-set the splined drive shaft 50 is rotatable in ball bearing ll supported in the end wall 56 and has integral therewith the flange l3 and ring gear 12. Ring gear 12 has internal helical gear teeth 14 and a series of spokes 16 extending radially from the outer periphery, the gear teeth 14 being the means through which power is applied during gear drive and the spokes 18 being the means through which power is applied during direct drive.

The driven shaft 18 of the underdrive gear-set is cut in one piece with the worm 88 of the axle gearing. A large double row ball bearing 82 is held in hub 84 between a shoulder 86 and the end of the cap 88. The cap 88 is secured by screws 98 and the bearing is secured to the worm by the nut 82. The bearing 82 carries half the radial load and all of the thrust of the axle gearing.

A smaller ball bearing 94 slidably held in hub 86 and free to move endwise therein carries the other half of the radial load of the axle gearing.

guideways I52 to and into the circular groove I50,

A small roller bearing 68 rotatable in the end of the drive shaft 50 supports the end of the driven shaft 18 but due to the nature of the underdrive gearing this bearing is required to carry substantially no load either radial or thrust.

The underdrive planet pinion carrier is made up of two main parts, the one part comprising an internally splined hub I00 snugly fitted over external splines I 02 on driven shaft 18 with an outwardly extending flange I04, and the other part comprising a drum I06 with an inwardly extending flange I08 and a large hollow hub I I0 extending integrally from the flange I08 to the flange I04 (see Fig. 20).

The hub IIO extends over the outside of the flange I04 at II2 to maintain concentric relation, and rivets I I4 hold the hub securely to the flange. At four equally spaced points (see Fig. 14) the huh I I0 is completely cut away as at II6 to make room for the planet pinions II8. Studs I20 are located centrally of the openings II6, the ends being held and riveted in the flanges I04 and I06 (see Fig. 1 or 2). The planet pinions II8 are rotatable on roller bearings I22 carried on the studs I20 and are thereby held in correct mesh with the internal teeth 14 of the ring gear 12.

The outside of the internally splined carrier hub I00 is ground smooth to provide a journal upon which the sun gear I24, shown in detail in Fig;

16, may rotate. A hearing bushing I26 is press fitted to the inside of the sun gear. An integral hub I28 extends rearwardly from the sun gear and is enlarged at I30 to provide a place for radial openings to contain the balls I32 and springs I34. A band I36 surrounds the hub to retain the springs. The extreme end of the hub is formed to compose jaw clutch teeth I38 which it will be noted are hooked as at I39 to maintain engagement under load. The gear teeth I are helical and at such an angle as to create an end thrust in the direction which urges the jaw teeth I38 to remain engaged.

concentrically secured by the screws I40 to the end of the hub 96 is the flanged jaw clutch member I42, shown in detail in Fig. 1'7. A hub I44 extending from the flange has cut integral thereon the jaw clutch teeth I46 which correspond to and are engageable with the jaw clutch teeth I38 of the sun gear I24.

A prolongation I48 of the hub I44 extends into the space left between the inside diameter of the sun gear and the smaller end I49 of the carrier hub I00.

Near the end of the hub I48 a round bottomed groove I50 extends completely around it. From this circular groove at equally spaced points around it other round bottomed grooves extend somewhat helically, forming the guideways I52 within which the balls I32 act as followers which may move to carry the sun gear I24 endwise along the hub I48. The guideways are somewhat deeper at their ends I54 than they are where they join the groove I50 so that radially inward pressure on the balls creates a slight tendency to cause the gear to move toward the jaw clutch member I42.

Fig. 2 shows the sun gear I24 when it is moved endwise as far as it will go with the jaw clutch teeth I38 fully meshed with the jaw clutch teeth I46 carried by the flanged member I42, and with the balls I32 in the deep 'ends I54 of the guideways I52. In this position the sun gear I24 is held against backward rotation as it must necessarily be held to provide gear drive. The sun gearmay, however, move axially on the hub I48 into the space I56 by drawing the balls I32 along the whereupon the sun gear is free to rotate forwardly as it must during direct drive.

The weight of the balls I32 and the strength of the springs I34 is preferably such that the centrifugal force of the balls becomes greater than the strength of the springs when the sun gear rotates about 600 R. P. M.

This proportion will allow ample pressure on the balls inasmuch as the only time the balls need become operative as followers to press downward in the guide-ways and guide the jaw clutch into engagement is when the sun gear I24 has completely ceased rotation and starts to rotate backwardly.

The balls I32, therefore, are never in frictional engagement with the groove I50 or the guideways I52 except a fraction of a second each time the change from direct drive to gear drive and vice verse. is taking place. As soon as the sun gear I24 rotates forwardly in direct drive the balls are drawn out of contact with the guideways and groove by centrifugal force. I

The guideways I52 are so located with respect to the teeth I46, and the balls I32 are so located with respect to the teeth I38 that whenever the balls follow the helical paths the mating clutch teeth approach each other in proper relation for full depth engagement. This is important, for where a jaw clutch is employed and is permitted -to engage without such guiding means, it frequently happens that the mating teeth engage with a very shallow hold thus throwing an exces- I sive strain on the points of the teeth which results usually in the engaged teeth slipping off and creating a jerk in the carrying of the-load. Such slipping off also soon breaks away the sharp corners of the teeth.

Contained within the drum portion I06 of the planet pinion carrier are the several parts which comprise the direct drive jaw clutch. A driven clutch ring I58 is secured to the drum I06 to rotate therewith but has limited axial movement therein. One face of the clutch ring I58 has jaw teeth I60 thereon (see Figs. 11 and 12).

An inner driving clutch ring I62 closely surrounds the ring gear 12 and has lugs I64 on one face fitting closely between the spokes 16 of the ring gear, and has jaw teeth I66 on the other face corresponding to and adapted for engagement with the jaw teeth I60 of the driven clutch ring I58. An outer driving clutch ring I68 surrounds the inner ring I62 and has lugs I10 on one face also extending between the spokes 16. The lugs I10, however, do not nearly fill the space between the spokes by an amount equal to the spaces I12 whereby the outer ring I68 is driven by the spokes 16 with considerable lost motion. The opposite face of the outer clutch ring I68 also has jaw teeth I14 which correspond to and are adapted for engagement with the jaw teeth I60 of the driven clutch ring I58.

The lost motion spaces I12 correspond in circumferential extent exactly to the circumferential, measurement of a jaw tooth I14. Therefore when the lost motion of the outer driving clutch ring I68 is taken up in one direction the jaw teeth I14 on its opposite face are misaligned with the jaw teeth I66 of the inner driving clutch ring I62, as in Figs. 11 and 14, but when the lost motion of the outer driving clutch ring I68 is taken up in the other direction its jaw teeth I14 will be aligned with the jaw teeth I66 of the inner driving clutch ring I62, as in Fig. 12. It will be seen that when the jaw teeth are arranged as in Fig. 12, the direct drive clutch may become engaged but when they are arranged as in Fig. 11, it may not become engaged.

A backing ring I16 is held in the drum I06 by the spring ring I18 which is sprung into a groove in the inside of the drum. The ends of the lugs I64 of the inner driving clutch ring I62 rest against the backing ring I16 thereby-preventing said ring being pushed out of the drum.

The lugs N0 of the outer driving clutch ring I68, however, are shorter than the inner ring lugs I64 and therefore do not reach the backing ring (see Figs.. 18, 11 and 12). A circular row of ground steel balls I is interposed between the shoulders of the rings I62 and I68 whereby the outer ring may be pushed against the inner, and the inner thereby be pushed against the backing ring. The balls I80 provide antifriction bearing means whereby the outer ring I68 may rotate with respect to the inner ring I62 through the lost motion space I12 with little resistance when the springs I84 have pressed the ends of the teeth I60 of the driven clutch ring I 58 (see Fig. 11) against the ends of the teeth I66 and I14 of the driving clutch rings I62 and I68 preparatory to effecting direct drivingclutch engagement. A shoulder I82 in the drum I06 prevents endwise movementof the clutch rings I62 and I68 in the drum, the said rings being fitted to rotate freely between this shoulder and the backing ring I16.

In the drawings the direct drive clutch above described is shown fully disengaged as it will always be whenever gear drive is in eifect. There is, however, means constantly urging the direct drive clutch into engagement. This means comprises the springs I84 contained in the openings' I86 of the hub IIO (see Figs. 14, 15 and 18).

Springs I84 act against the heads I88 of the studs I90 whereby the pressure plate I92 is caused to push the pressure pins I94 against the driven clutch ring I58 thereby urging said ring always toward the driving clutch rings I62 and I68. The pressure pins I94 extend slidably through holes in the flange I08 and are then reduced in diameter at I96 where they enter the driven clutch ring I58. Pins I94 therefore not only act to push the ring I58 axially but also compel rotation of the ring in unison with the drum I06.

The speed responsive mechanism which is provided to vary the stress of the springs I 8*. and thereby vary the clutch engaging pressure comprises the weights I98 hingedly supported by pins 200 extending through ears 202 which are carried on the face of the flange I08. Inwardly of the hinge pins the weights are shaped in the form of a cam 204 against the heel of which the spring stressing plate 206 normallyrests (see Fig. 19). Guide pins 208 extend from plate 206 into holes in the hub IIO to insure that no weight I98 may move outwardly from the axis faster than the others and thereby create an unbalanced effect. A ball thrust bearing 2| 0 is interposed between the enlarged portion I30 of the sun gear and the pressure plate I92 whereby axial movement of the sun gear into the gear drive position shown in the drawings forces the pressure plate back against the stress of the springs I84 to completely disengage and prevent rubbing between the driving and driven rings of the direct drive clutch, the helix angle of the teeth I4I of the sun gear and its hooked jaw teeth I38 both cooperating to maintain this relation as long as the sun gear is under suflicient load.

The extent .to .which the sun gear I24 must be unloaded during gear drive to eflect direct drive, I

and the extent to which it must be overloaded during direct drive to effect gear drive is determined by the position of the weights I 98 and therefore by the vehicle speed.

Fig. 21 shows diagrammatically the operation of the centrifugal weights I98 and their cams 204 on the spring stressing plate 206 and the springs I84. The diagram is drawn to a scale double that in the remainder of the drawings.

Assuming for illustration that the mechanism herein described is to be used with a 90 H. P. engine capable of delivering a maximum of foot pounds torque, the eight weights I98 should together weigh 1.67 pounds. The diagram Fig. 21 assumes that the eight weights have been combined into one and that its center of gravity when the weight is at rest and as close to the axis of rotation as it may get will be at a and that when the weight has swung out as far from the axis of rotation as it may, its center of gravity will be at "9", the points "1), 0', "d, e, and

"i" being intermediate positions of the weight.

The point It represents a hinge pin 200. The line i represents the axis of rotation. A line drawn between a and h is at an angle of 30 degrees with the axis 2' and a line is drawn between "g and h is at an angle of seventy-five degrees with the axis i. Points b, "c, d, e and "f are equally spaced between points "a llgli.

The line 1'. represents the rear face of the spring stressing plate 206 against which the cam 204 rests when the weight is at "a". The lines k, Zn, am, a (0)) and up represent positions to which the rear face of the spring stressing plate 206 is moved by the cam 204 when the weight takes the positions "1), c, d, e, f and g respectively.

It will be observed that, when the weight is in the home position a. (see also Fig. 19) the heel of the cam 204 bears on the rear face of the spring stressing plate 206, but when the weight is -in the extreme outward position "9", whereby the rear face of the spring stressing plate has been moved to the line "1), the toe of the cam 204 bears on the rear face of the spring stressing plate 206.

In the intermediate positions b, c, d e and f of the weight whereby the plate has been moved respectively to the lines k, l, m, "n and 0" the cam bears on the plate at points intermediate the heel and toe.

By reference to the diagram Fig. 21 it will be seen that, when the weight is at a it is 4.250 inches from the axis of rotation i and the centrifugal force which this 1.67 pound weight exerts at this distance from the axis is applied to the back of the spring stressing plate 206 through a leverage of 1.875/ .375, but that when the weight is at g, the centrifugal force which the 1.67 pound weight exerts at 5.256 inches from the axis is applied through a leverage of .561/.990, and that at intermediate positions of the weight the force it applies to the plate is applied through intermediate leverages.

By employing eight springs of round wire coiled pitch diameter, having a lead of about 4 turns per inch, and having a free height of 2.238, and by having these springs already compressed to 1.875" when the weight'is in the position "a, it may be found by simple mathematics that the springs will provide an initial stress of 229 pounds (see Fig. 21) that the weightin assuming the positions b, .c d,

"e, "1 and a acting through the cam will shorten the springs to lengths 1.815, 1.745, 1.650, 1.550, 1.438, and 1.313" respectively. a

At these lengths the spring stress will be 267, 310, 371, 434, 505 and 584 pounds respectively, and the weight, to equal these stresses when in the positions ab, a a v, n w 4:! and an, must, when acting through the progressively decreasing leverage indicated, he revolving 609, 757, 951, 1186, 1517 and 2040 R. P. M. respectively.

By reference to Figs. 11 and 12, it will be seen that the driving teeth I66 and Ill oi the direct drive clutch are beveled about 45 degrees. There is, therefore, always as much force urging the driven clutch ring I58 axially out of engagement as the tangential load carried by its teeth I60. It follows that the tangential load which the direct drive clutch will carry, without being forced out of engagement thereby, varies as the stress of the springs which varies as the R. P. M. of the weights which in turn vary as the vehicle speed.

In the chart Fig. 22, the vertical column of figures at the left indicates the foot pounds torque which it is possible for the engine to deliver, the column at the right, the thrust which said foot pounds create in trying to disengage the direct drive clutch in opposition to the clutch engaging effort of the springs, and the figures at the bottom the M. P. H., of the vehicle and cor--- responding R. P. M. of the weights.

The curve q is plotted in accordance with the two columns of figures in Fig. 21 designated pounds stress and R. P. M. of weights. Thus the curve passes through the points where 2040 R. P. M. intersects 584 pounds stress, where 1517 R. P. M. intersects 505 pounds stress, 1186 R. P. M. intersects 434 pounds stress, etc.

By this curve it may be seen that, at 10 M. P. H., there may be applied 77 foot pounds engine torque, or 45% of maximum engine torque without forcing the direct drive clutch out of engagement and causing gear drive; that at 20 M. P. H., there may be applied 127 foot pounds or 75% of maximum without disengaging the clutch; and that at 30 M. P. H., 162 foot pounds or 95% of maximum torque may be applied.

In order to more clearly show the reason for progressively decreasing the leverage through which the weight acts as it moves farther from the axis of rotation as indicated in Fig. 21, a curve "1" is drawn showing how a conventional centrifugal weight having capacity to produce 584 pounds spring stress at 2040 R. P. M. falls off rapidly in capacity at the lower speeds.

By referring to the curve "r it will be seen that at 10 M. P. H., there could be applied only 15 foot pounds, or 9% of maximum engine torque without forcing the direct drive clutch out of engagement and causing gear drive; that at M. P. H., there could be applied only 58 foot pounds or 34% of maximum engine torque without disengaging the clutch, and that at 30 M. P. H., 133 foot pounds or 78% of maximum of maximum engine torque may be applied. A line s is drawn on which the pounds thrust is in direct proportion to the R. P. M. to show by comparison the difference between the thrust developed by the herein described clutch engaging mechanism and the conventional. It will be seen that on the curve 1' at the lower speeds the thrust increases only in direct proportion to the R. P. M., but at the higher speeds it increases substantially in proportion as the square root of the R. P. M.

It is obvious, however, that the active surface of the cam 204 may be shaped so as to produce any rate of increase in thrust, in proportion to the speed, within reason.

The foregoing comparison makes clear one of the reasons why speed-torque transmissions have not to date been commercially adopted, speed torque mechanisms being those wherein the direct drive clutch is held in engagement, by a speed responsive device and urged out of engagement by a torque means, because when weights were designed to properly balance the torque at high speed they created forces wholly insufficient at the lower speeds.

The result was that any attempt to drive in direct drive at the lower speeds resulted in throwing the mechanism into gear unless very high engine torque was created and applied.

The curve t shows the maximum engine torque which may be applied at different vehicle speeds. It will be observed that this curve is substantially flat between 5 and 35 M. P. H.

Now, such a curve could not be had where the engine was coupled to the vehicle through a friction clutch which permitted but little slippage. But where a fluid coupling is employed it will be found that, by the time the vehicle has reached a speed of 5 M. P. H., and the driven element 44 of the fluid coupling, acting through the underdrive gear-set, is revolving about 490 R. P. M., the driving element 34 and therefore the engine, have speeded up through slippage to about 1000 R. P. M., at which point the engine is delivering its maximum torque.

Inasmuch as a planetary gear-set such as is herein shown becomes structurally impractical if made with a ratio greater than about 1.75 to 1, the combination of such a gear-set with a fluid coupling is important, for, where the engine is connected to the gear-set without slippage, the engine cannot, at the lower vehicle speeds, run ahead of its maximum torque point, and therefore, where the engine is thus solidly connected to the gear-set, there must be provided a gearset giving considerably greater reduction to enable the engine to accelerate the vehicle from the lower speeds with sufficient rapidity. When in turn a gear-set having such greater reduction is provided, there results considerable engine racing when the vehicle is being accelerated up to about 35 M. P. H., through the gears.

In the combination herein disclosed the gearset has the ratio of about 1.625 to 1. This ratio becomes ample for acceleration at low vehicle speed because the engine may race ahead to its highest torque point, and it does not provide too great a reduction at the higher speeds to cause engine racing because, at the higher speeds, there is substantially no slippage in the fluid coupling and the engine therefore runs no faster than the driven fluid coupling member 44.

Due to the great angle of the helical teeth III of the sun gear I 24 and the hooked-under nature of its jaw teeth I38, and the further fact that the point of application of the tangential load on the sun gear is closer to the axis than the torque load applied to the direct drive clutch teeth I60, it may be found that the engine torque required to force the sun gear to its extreme operative position shown in the drawings is about half the torque necessarily applied to force the ,direct drive clutch out of engagement.

The curve 11. is plotted with half the value for a given speed as the curve "q. From the diagram it will be seen that, having created any torque curve similar to T which, at the existing speed, is somewhere above the curve "q and thereby enforced disengagement of the direct drive clutch and engagement of the gear drive, it will be necessary to reduce the torque t until it is somewhere below the curve u before the sun gear moves toward the space I56 farenough to permit the faces of the teeth I66, (see Fig. 11)

to rub against the faces of the teeth I66 and I14. After this the direct drive may best be re-established by releasing the accelerator and allowing the braking action due to the friction between the faces of the teeth I66 and the teeth I66 and I14 to aid engine deceleration until synchronism between the said teeth is established.

During gear drive the engine is always rotating the driving clutch ring I68 faster than the speed of rotation of the driven clutch ring I58. When the applied engine torque is reduced to a point below the curve 2!. and the faces of the direct drive clutch teeth come together as above indicated, the outer clutch ring I68 is dragged backward in the direction of the arrow 2I2 by the frictional contact until its lugs I16 touch the spokes 16 whereby the teeth I66 and the teeth I14 are misaligned thereby preventing entry of the teeth I66 as long as the speed of the driving clutch rings is faster than that of the driven.

When, however, the engine, aided by the friction between the ends of the direct drive clutch teeth, loses suflicient momentum to establish synchronous speeds between the driving and driven clutch rings, any further decrease of engine speed which is sufllcient to cause it to lose "3 2' of a revolution with respect to the driven clutch ring, causes the outer driving clutch ring I68 to be dragged forward in the direction of the arrow 2 I4 until its lugs I 16 engage the spokes 16, whereupon the driving clutch teeth I66 and I14 are aligned as in Fig. 12 and the driven teeth I66 will be seated between the driving teeth I66 and I14 by the force of the springs I84. It is obvious that the direct drive clutch will thus engage without claslr. The driven shaft 18, and consequently the worm 86, will be revolved either at engine speed or at 1/ 1.625 or 61 of engine speed, depending on whether the direct or gear drive is in effect, assuming, in either case, that slippage between the driving and driven elements of the fluid coupling has substantially ceased. In starting from a dead stop, however, it may be expected that the worm 86 will not, at speeds much below 5 M. P. H., be revolving more than 25% of engine speed.

Within the axle housing 64 the worm 86 is in constant mesh with the hollow worm wheel 2I6. Worm wheel 2I6 has end supporting flanges 2I8 and 226 secured to the worm by screws 222 and 224 respectively. Hubs 226 and 228, carried respectively in the flanges 2I8 and 226, are rotatably supported in ball bearings 236'carried in the bearing plate 68 and end head 16.-

The advantage of employing a worm and wheel as axle gearing, instead of the more commonly employed bevel gearing, in combination with the fluid coupling arranged as shown will be apparent when it is considered that, if bevel axle gearing were used in such a combination the center line 2--2 of Fig. 1 would coincide with the center line 3-3 thereof. The bottom of the flywheel would then be so near the road surface as to make the whole structure much less desirable.

Immediately inside the hollow worm wheel is the overdrive planet pinion carrier which is also a hollow structure comprising the member 232 having end heads 234 and 286 secured thereto by screws 288 and 246 respectively. Hubs 242 and 244 are carried centrally in the end heads 224 and 236 respectively. The hub 242 has rotative bearing on a short bronze bushing 246 carried on the outside of the hub 226, while the hub 244 has rotative bearing in a bronze bushing 248 carried within the hub 228. The bronze bushings 246 and 248 are operative only when the vehicle is being driven backwardly.

The inwardly extending portion of the hub 242 is cut away at four places 256 (see Fig. 6) for the overdrive planet pinions 252. Supported in the hub 242 at opposite sides of the slots 256 are the pinion studs 254 which carry the roller bearings 256 upon which the pinions rotate.

A sun gear 258 having teeth 256 in constant mesh with the teeth of the pinions 252 has a long hub 266 supported in a bronze bushing 262 held within the hub 226. The sun gear carries no radial load and the bronze bushing is therefore adequate to hold it centrally supported.

The differential pinion carrier 264 contains the bevel pinions 266 rotatable upon the spider 268. The bevel gears 216 are integral with the axle shafts 212 and 214. A flange 216 extending from the carrier 264 has the ring gear 218 secured thereto by the screws 286, the teeth 28I of the ring gear being in constant mesh with those of the planet pinions 252. The differential carrier 264 is rotatable in ball bearings 282 one of which is supported directly in the end of the member 232 and the other in a recessed end of the hub 242.

In a planetary gear-set of the type shown having as main elements in addition to the planet pinions, a ring gear usually designated as R, a planet pinion carrier designated as C, and a sun gear designated as S, it is well known that (1) if S is held against rotation, R" is made the driver, and C the driven, as is done in the case of the underdrive gear-set hereinbefore described, a reduction in speed will be provided; (2) that if 8" is held against rotation, C is made the driver, and R the driven, an increase in speed will be provided; (3) that if any two members such for instance as "S and C are both made drivers while R is made the driven a direct drive will be provided; (4) that if C is held against rotation while S is made the driver and R the driven, "R will rotate in the reverse direction; and (5) that if S only is made the driver while R" is connected to the driven elements and C is left wholly free, "0 will revolve idly, slowly forward, and no driving connection will be had between the driving and driven members.

The underdrive gear-set first described herein operates by making the first of the above connections, while the overdrive gear-set last described has means for making connections 2-5, manual control means being provided to elect between operating the vehicle forwardly, operating it rearwardly, or letting it stand still while the engine rotates, and automatic means being provided to change from direct forward to overdrive forward and vice versaat predetermined vehicle speeds.

The manual control means is contained in the end head 16 and comprises the sliding collar 284 having internal splines fitting slidably over the external splines 286 of the hub 244, internal clutch teeth 288 slidable over externalteeth 286 formed on the hub 228, and external clutch teeth 282 slidable into the internal teeth 284 of the plate 68, whereby said ring is non-rotatable but axially 236 which is held against rotation to the end head 10 by the rivets 298.

A fork carries rollers 302 on studs 303 which extend into a groove 305 in the collar 284. A splined shaft 304 extends from the lever 306 through a splined opening in the fork. A nut 308 holds the shaft in place but leaves it free to rotate. A detent ball M0 and spring 3I2 (see Fig. 10) retain the lever 306 in its several positions. Any convenient connecting means (none shown) may be provided whereby the operator may move the lever 306 to its several positions. A lid 3l4 held on by screws 3I6 facilitates assembly of the fork in the end head.

The automatic control means is contained in the head 66. Its function is to normally compel the sun gear 258 to rotate in unison with the worm wheel 2l6 to provide direct drive but to hold the sun gear against rotation whenever the applied power is momentarily released while the vehicle is operating above a predetermined speed to provide overdrive. To perform this function, jaw clutch mechanism is employed wherein one jaw clutch engages before the other disengages in order that there may be no free-wheeling at any time, not even in the transition period between direct and overdrive.

A" spider 3! has internal splines extending snugly into the external splines 320 of the axle shaft 212. Pairs of arms 322 extend outwardly for hingedly supporting the centrifugal weights 324 on the hinge pins 326. Stops 328 extend from the arms 322 and stop pins 330 in weights 324 limit outward movement of the weights. A collar 332 has axially extending driving lugs 334 ex tending between pairs of arms 322 whereby said collar is rotated in unison with the spider, but is axially movable with respect thereto. Lever arms 336 depending from the weights 324 extend into notches 338 in the ends of the lugs. The driving lugs 334 are long enough that they do not come completely out of the spaces between the arms 322 when the weights 324 are in their extreme outer position.

The axially operable clutch member 340 has internal splines slidably fitted over the external splines 342 on the long hub 260 of the sun gear 258. The collar 332 surrounds the smaller end of the member 340 and shoulders against said member at 344. Axial movement of the collar 332 by the weights 324 therefore moves the clutch member 340 axially, although the said collar and clutch members are rotating at different speeds.

Driving the centrifugal mechanism in unison with an axle shaft insures that the transition between direct and overdrive and vice versa will always be determined by the vehicle speed independently of what engine speed is being used to maintain it. A spring 346 is interposed between the end of the bushing 262 and the hub of the clutch member 340 whereby the said clutch member 340, collar 332 and weights 324 are normally held in the positions shown. The end of the clutch member 340 has external clutch teeth 348 and internal clutch teeth 350.

Axially slidable on the external splines 352 of the hub 226 is the internally splined clutch hub 354 which has external clutch teeth 356 normally engaged, as shown, with the internal clutch teeth 350 of the clutch member 340. An integral flange 358 extends outwardly from the outside of the hub 354.

A clutch ring 360 has short external splines axially slidable in the somewhat longer internal splines 362 formed on the interior of the head movable. A hub on one side of the ring 360 has internal clutch teeth 366'.adapted to receive the external-clutch teeth 348 of the clutch member 340. An integral flange 368 extends inwardly from the body of the ring 360. A bronze wear washer 310 is interposed between the flanges 358 and 388.

The hub of the clutch member 340 has six equally spaced holes 3 extending axially therethrough, three of them containing springs 312 and the other three containing studs 314. A ring 316 has six equally spaced bosses 318 over three of which the ends of the springs extend, and into the other three of which the small ends of the studs 314 are riveted.

The heads of the studs 314 limit expansion of the springs to the length shown except when the clutch member 340 is moved axially by outward movement of the weights 324. A washer 382 has internal splines extending into the external splines 342 to insure that the base upon which the ends of the springs 312 rest will always revolve at the same speed as the holes in the clutch member which contain the springs.

At three circumferentially equally spaced -points in the face of the clutch ring 380, are

pockets for holding the one end of the springs 384, the other end being held in correspondingly spaced pockets in the bearing plate 68. The springs 312 and 384 are exactly alike but are here given different numerals to facilitate description. A shoulder 364 limits end movement of the ring 360 by the spring 384.

In the drawings the axle mechanism (see Fig. 3) is shown as it would appear when at rest. The weights 324 are in their innermost position. The internal clutch teeth 350 of the clutch member 340 are in full engagementwith the external clutch teeth 356 of the clutch hub 354. The clutch member 340 being splinedly mounted on the hub of the sun gear, and the clutch hub 354 being splinedly mounted on the hub of the worm wheel 2H5, it follows that the only effect of the entire automatic clutch operating mechanism jacent meshing faces of the clutch teeth of the directto over-drive connecting mechanism are beveled to insure transition from direct to overdrive and vice versa without an interval between wherein there is no connection between the engine and vehicle wheels.

The two middle rows of teeth 348 and 350 are actually carried on the outside and inside respectively of the clutch member 340, as in Fig. 3, but are shown in Fig. 9 for illustrative purposes as being side by side, the teeth 356 being on the clutch hub 354 and the teeth 366 in the nonrotatable ring 360. Normally the teeth 358 and 350 are fully engaged, as in Fig. 3, for direct drive. the teeth 348 and 366 being fully meshed only after the transition period to overdrive.

The faces 343 are beveled so that the edges 35! of the teeth are only A as Wide as originally. Thus the teeth 356 may enter into the spaces between the teeth 350 until the edge 353 extends over the teeth 363 as long as the teeth 348 do not rotate faster than the teeth 355. In practice, as in Fig. 3, the abutting flanges 358 and 358 limit the distance that both sets of teeth may be en-- tered, at the same time, to about one-third full depth, instead of one-half full depth. The manher in which this arrangement operates will be hereinafter more fully described.

Proportion While the transmission shown may be designed for an engine of any horsepower, some indication of its proportion for a given horsepower may pref erably be given.

With an engine of 85 to 90 H. P., at 3800to 4200 R. P. M. and a total vehicle weight of 2600 to 2900 pounds, the proportion of most of the parts may be gotten by taking the largest diameter of the worm wheel 2i5 as 9." and making all other parts of the mechanism to the same scale. Some of the dimensions which are not readily gotten by sealing the drawings are as follows:

The helix angle of the underdrive gear-set should be 45 degrees. The ring gear should have a pitch diameter of 5.6568" and have 64 teeth; the sun gear a pitch diameter 3.5355" and have 40 teeth; and the planet pinions a pitch diameter 1.0606" and have 12 teeth. The rule for ratio when the sun gear is held against rotation and the ring gear is the driver is, one revolution of the driven carrier C is produced by revolutions of the driver R. The underdrive ring gear R must therefore revolve revolutions of the carrier C.

In planetary gearing of the type herein employed the ratio available is confined within narrow limits, being always less than 2 to 1 and more than 1 to 1, the practical limit being reached at about 1.75 to 1 for most and 1.25 to 1 for least reduction. A ratio of 2 to 1 would be obtainable only if it were possible to make the diameter of the sun gear and ring gear equal, which would make the planet pinions zero diameter, while the ratio of 1 to 1 would be obtainable only were it possible to make the planetshalf the ring gear diameter which would require a sun gear of zero diameter. The hook under angle I39 of the sun gear clutch teeth I38 is preferably about 30 degrees.

The proportion of the centrifugal weights I98 and of the clutch engaging springs I84 of the underdrive gear-set have been hereinbefore given.

The worm 80 has a pitch diameter of 1 and a quintuple thread the lead angle of which is 45 degrees. The worm wheel 2l5 has a pitch diameter of 8%" and has 25 teeth. The worm and wheel therefore has a ratio of 5 to 1.

The helix angle of the overdrive gearing is 23 degrees. The ring gear has a pitch diameter of 5.431" and has 40 teeth, the sun gear a pitch diameter of 2.172" and has 16 teeth, and the planet pinions a pitch diameter of 1.629" and have 12 teeth.

When reversing is to be done with gearing of this type the sun gear is made the driver and the carrier is held against rotation, the ring gear being the driven member. The rule in this case is 1 revolution of the sun gear produces S/R revolutions of R backwardly. The reversing ratio is then 16/40=%. That is, one revolution of the sun gear produces of a revolution of the ring gear backwardly.

When overdrive is to be effected, the sun gear is held against rotation and the planet pinion carrier is made the driver. The rule in this case is 1 revolution of the carrier produces revolutions of the ring gear. The overdrive-ratio is then that is, 1 revolution of the driver carrier produces and heavy load 8%; to 1; for direct drive with moderate load and. speeds to 50 M. P. H., to 1; for overdrive, moderate load and speeds over 50 P. H., 3.57 to 1; and for reversing, ordinary load, 12% to 1, except when the reversing pull is very heavy and thereby brings the underdrive gear-set into action whereupon the reversing ratio will be 20 to 1.

In common practice the 8% to 1 for lowest available forward engine-to-wheel ratio would be considered insufficient, a ratio of 10 or 11 being more often used. But the 8% to 1 reduction, when combined with a fluid coupling will give better results than a 10 to 1 ratio employed with the conventional clutch, because of the fact with the fluid coupling the engine may run ahead and reach its most efiicient torque point while the vehicle is still moving at the lower speeds.

Also, in common practice, the 5 to 1 engine-towheel ratio for direct drive would be considered too slow, a ratio of 4 to 1 being more often used. But this is not because the 5 to 1 ratio would not be better at speeds below 50 M. P. H., but because vit ,would result in too much engine rushing at speeds around 70 M. P. H. Where, however, a ratio of 3.57 to 1 is provided in the form of an overdrive for speeds above 50 M. P; H., the 5 to 1 direct drive ratio for speeds under 50 M. P. H., is much more desirable.

Operation The operation of the mechanism may .be carried out as follows:

With the manually shiftable lever 306 in the neutral position shown the engine may be started and rotated for warming it up when required. The sun gear 258 is now connected to the worm wheel 2l6 through the automatic mechanism within the head 65. The ring gear 218 being held stationary by the resistance which the vehicle offers to being moved, causes the planet pinion carrier 232, which is not now connected to anything, to revolve slowly forward at 2/35 engine speed. r

driven backwardly at 12 /2 to 1 engine-to-wheel ratio, except under severe load in which case the underdrive gear-set will cut in and the reversing ratio will be 20 to 1.

The manually shiftable lever 306 may next be shifted to cause engagement of the clutch teeth v 288 with the clutch teeth 290. This connects the planet pinion carrier 232 to the worm wheel 216.

The sun gear 258 is already connected to the Worm wheel 2H5 through the automatic mechanism in the head 66.

In this condition the planet pinions 252 cannot rotate on their studs 254 and therefore the ring gear 218 is also driven forwardly at the same speed as the worm wheel 2|6, whereby there is provided, for forward driving, an engine-to-wheel ratio of 5 to 1, unless an engine torque somewhere above the curve qf, Fig. 22, is applied, in which case an engine-to-wheel ratio of 8 to l is provided by cutting in the underdrive gearset.

With the manual connection thus made for forward driving this connection may be maintained unless and until it is desired to race the engine or reverse the vehicle. While this connection is in effect the vehicle may be started from rest and operated with the 5 to 1 ratio, if rapid acceleration is not desired, by keeping the applied torque between the curves q and u Fig. 22. The 8 to 1 ratio may be brought in at any time below 34 M. P. H., by momentarily applying torque above the curve q after which this 8%; to 1 will remain in effect until the torque is reduced below the curve u whereupon the 5 to 1 ratio may be reestablished. Once established, the 8 to 1 ratio may be continued as long as desired, but may be exchanged for the '5 to 1 ratio at any time by momentary release of the accelerator pedal.

The foregoing connections cover all ordinary driving, but at speeds over 50 M. P. H., the overdrive gear-set may be brought into play to reduce the engine speed. This is also done by momentary release of the accelerator pedal after a speed of 50 M. P. H., is passed.

Below 50 M; P. H., the force of the large spring 346 alone is sufiicient to hold the weights 324 in the position shown. Above 50 M. P. H., the spring 346 plus the friction, under load, of the engaged clutch teeth 350 and 356 will hold the weights to the position shown.

But any time, after a speed of 50 M. P. H. has been passed, that the accelerator pedal is released, and the friction, under load, of the teeth 350 and 356 is reduced to near zero, the force of the spring 346 will be insufficient to hold the weights 324 and they will quickly move outward until the stop pins 330 are arrestedby the stops 328.

As the weights 324 move outwardly the clutch member 340 moves axially whereupon the faces of the teeth 348 engage the faces of the teeth 366, but inasmuch as these teeth cannot engage because the clutch member 340 is revolving about 600 R. P. M. and the ring 360 is non-rotatable, the ring 380 is pushed axially up to the bearing plate 68 against the resistance of the light springs 384.

Now this axial movement of the ring 360 1 when the 'end of the hub 354 encounters the edge 355 of the ball bearing 230, as it does after movement, the flange 358 can move no further and the clutch member 340 and ring 360 continue moving on until the flange 368 is drawn away from the flange 358 leaving a gap between them of and the teeth 350 are pushed twothirds'of the way out of mesh with the teeth 356 leaving them meshed or one-third of their width.

Now it will be remembered that the fuel was momentarily interrupted to allow the weights 324 to move out and re-arrange the clutch members as above stated, and if this interruption is allowed to continue for about three seconds, that is, until the engine and worm wheel have been reduced to of their former speed, while the ring gear remains at substantially the same speed due to vehicle momentum, the sun gear will thereby have been caused to lose enough speed to bring it to zero revolutions.

The reason why the sun gear may lose this much speed with respect to the worm wheel when they are directly connected by the onethird mesh of the teeth 35!] and 356 is because of the way their adjacent faces are beveled at 349 as explained relative to Fig. 9, that is, the teeth 350 which revolve in unison with the sun gear 258 and the teeth 356 which revolve in unison with the worm wheel 2 l6 are beveled in such a manner that the sun gear may lose speed with respect to the worm wheel by the ratchet effect of the beveled faces 349 when the teeth are onethird meshed but may not gain speed in respect therewith, loss in speed being a movement in reverse of the arrow 380.

.Now to maintain direct drive connection it is only necessary to keep the sun gear from rotating at a speed in excess of that of the worm wheel, and, inasmuch as the ratchet effect of the said beveled faces prevents such excess speed, it follows that the direct drive connection is not unmade while the sun gear is being brought to zero revolutions preparatory to holding it against rotation for overdrive.

Thus if the operator waits only one second after he has interrupted the fuel and the weights have moved out and rearranged the clutch parts as above described, and he then reapplies the fuel before waiting the other two seconds required to bring the sun gear back to zero revolutions to make full depth overdrive connection, he merely operates in direct drive, although the weights are out, because the ratchet action of the onethird engagement of the teeth 35!) and 356 prevents the speed of the sun gear exceeding that of the worm wheel and consequently maintains direct drive.

If, however, after the first second has elapsed, and the weights are out,-and the clutch parts have been rearranged as above indicated, the operator waits the additional two seconds until the sun gear loses it momentum and reaches zero revolutions the springs 384 will expand and move the non-rotatable clutch ring 360 away from the plate 68 and fully mesh its teeth 366 with the now non-rotating teeth 348 which are connected to revolve in unison with the sun gear.

As the ring teeth 366 slide over the teeth 348 -the fiange368 acts against the wear ring 318 andflange 358 and completely unmeshes the onethird meshed teeth 358 and 356 so that their beveled faces 349 need not ratchet over each other during overdrive operation. As long as the sun gear is held against rotation overdrive will be in effect.

Due to the fact that the weights 324 have more force at the same speed when in the out position than when in the in position, it is necessary to bring the vehicle speed down to about 45 M. P. H., before the spring 346 has force enough to move the weights back inward against their outward centrifugal force.

Any time then, at a speed below 45 M. P. H., that the operator momentarily interrupts the fuel of the engine, the reverse of the process hereinbefore described takes place. The spring 346 expands, moving the clutch member 340 axially, the teeth 358 of which shove the teeth 356 of the clutch hub 354 ahead against the resistance of the light springs 312 because the teeth 35!] are now non-rotating and the teeth 356 are rotating. The ring 368 follows until stopped by the shoulder 364. The member 340 continues on thereby pulling the flanges 358 and 368 apart, as before.

The ring teeth 366 and clutch member teeth 348 are left one-third meshed but their faces being beveled as at 349, Fig. 9, the teeth 348 may ratchet over the one-third meshed teeth 366 when power is reapplied.

The reapplied power causes the stationary teeth 348 to take up speed until they equal the speed of the worm gear 2 l 6, as the reapplied fuel causes the worm, which during overdrive is revolving at seven-tenths normal speed, to resume normal speed. When the sun gear gains sufficient speed to equal the speed of the worm wheel, the springs 312 acting against the collar 316 push the teeth 356 into the spaces between the teeth 358, and in doing so cause the flange 358 to act on the flange 368 and thereby unmesh the one-third meshed teeth 348 of the clutch member, from the teeth 366 of the ring.

By keeping the teeth 348 and 366 meshed onethird of their width until the direct drive teeth become meshed, no free wheeling is possible at any time. If the operator, below speeds of 45 M. P. H., interrupts the fuel and allows the weights 324 to go in, and he then fails to reapply the fuel and thereby fails to complete the meshing of the direct drive teeth 350 and 356, the engine will be driven by vehicle momentum as soon as the teeth 348 try to turn backwardly of the direction of the arrow 380 and encounter the teeth 366 with which they are meshed one-third of their depth.

From the foregoing it will be seen that the mechanism herein shown, by employing a relatively slow axle reduction having an engine-towheel ratio of 5 to 1 in combination with a fluid coupling, a speed range from to 60 M. P. H., may satisfactorily be had without other than axle gearing except when rapid acceleration is desired; that rapid acceleration may be had through gearing giving 8 to 1 engine-to-wheel reduction by momentarily applying a considerable proportion of the full engine torque, said proportion varying with the vehicle speed, less torque bringing in the gear at the lower vehicle speeds Where gearing is more necessary for acceleration; that when combined with a fluid coupling an 8 to 1 engine-to-wheel reduction is more effective at the lower speeds than a 10 or 11 to 1 reduction when connected directly to the engine; that the to 1 axle gearing is not too slow but preferable when an overdrive giving a 3.57 to 1 engine-to-wheel reduction is provided for speeds above 50 M. P. H.; that the changes from underdrive to direct, and from direct to overdrive are made without free wheeling either before, after, or during the transition period; that there are no spring, roller, or friction clutches contained in the mechanism; and that the number of different speeds provided are had with a minimum of gearing.

While the embodiment disclosed contemplates its use as a rear axle, the same may be used as a front axle by changing the worm from right to left, or the underdrive gear-set, the fluid coupling, and the engine may be more widelyspaced from the ,axle than in the embodiment shown, that is, more as in common practice.

The claims herein are all confined to the underdrive gear mechanism-adjacent the fluid coupling which differs from any heretofore known in that it is a speed torque controlled gear-set in which there are no friction clutches.

Having described my invention, I claim:-

1. Automotive speed reducing mechanism comprising, a driving member, a driven member, an

internal ring gear on the driving member, a

planet pinion carrier on the driven member, planet pinions carried by said carrier in constant mesh with said ring gear, a sun gear in constant mesh with said planet pinions, a jaw clutch normally connecting said ring gear and said carrier for rotation in unison, stressed resilient means holding said jaw clutch engaged, torque means to overcome said resilient means and means for holding said sun gear against rotation upon disengagement of said jaw clutch.

2. The structure defined in claim 1 with beveled driving faces on the teeth of the said jaw clutch constantly urging disengagement thereof upon overload.

3. The structure defined in claim 1 with beveled driving faces on the teeth of the said jaw clutch constantly urging disengagement thereof upon overload, and with speed responsive means constantly varying the stress of the resilient means as the speed of the driven member varies.

4. The structure defined in claim 1 .with beveled driving faces on the teeth of the said jaw clutch constantly urging disengagement thereof upon overload, and with a centrifugally operated means for varying the stress of the resilient means, and means operated by said centrifugal means which increases the stress at a rate which is less than in direct proportion to the rate of increase in the speed of the driven-member.

5. Automotive transmission mechanism comprising, a driving member, a driven member, a ring gear on the driving member, a planet pinion carrier on the driven member, planet pinions carried by said carrier in constant mesh with said ring gear, a sun gear in constant mesh with said planet pinions, a duodirectional direct drive jaw clutch normally connecting said ring gear and said carrier for rotation in unison, and a unidirectional gear drive jaw clutch for holding said sun gear against backward rotation upon disengagement of said direct drive jaw clutch.

6. The structure defined in claim 5 with a resilient means constantly urging engagement of the direct drive clutch and disengagement of the gear drive clutch.

7, Power transmission mechanism comprising,

, a driving member, a driven member, a jaw clutch for connecting said members directly, resilient means for holding the jaws of said clutch engaged, speed responsive means for varying said resilient means, torque responsive means for forcing said jaws into disengaged position upon overload, and gearing for connecting said members around the clutch upon disengagement thereof.

8. Power transmission mechanism comprising, a driving member, a driven member, a jaw clutch for connecting said members directly, stressed resilient means for holding the jaws of said clutch in engagement, torque responsive means urging the jaws of said clutch out of engagement upon overload, gearing for connecting the said members around the clutch upon disengagement thereof, and speed responsive means increasing the stress of said resilient means at a rate which is less than directly proportional to the increase in speed.

9. Power transmission mechanism comprising, a driving member, a driven member, a jaw clutch for rotatably connecting said members, resilient means holding the jaws of said clutch in engagement, torque responsive means urging said jaws out of engagement upon overload, and means for preventing reengagement of said jaws after said driving member has gained on said driven member such portion of a revolution as is represented by the circumferential width of one jaw, and for permitting reengagement of said jaws when said driving member has lost on said driven member such portion of a revolution as is represented by the circumferential width of one jaw.

10. Power transmission mechanism comprising, a driving member, a driven member, a jaw clutch for rotatably connecting said members, stressed resilient means for holding thejaws of said clutch in engagement, torque responsive means urging the jaws of said clutch out of engagement upon overload, a centrifugal weight movable away from the axis of rotation, and a cam operated by said centrifugal weight for increasing the stress of said resilient means, said cam being so shaped as to increase the said stress at a rate which is less than in direct proportion to the increase in the speed of rotation.

11. Power transmission mechanism comprising, a driving jaw clutch ring having teeth on one face, a second driving jaw clutch ring surrounding the first and having a like number of teeth on the same face, means to permit the second said ring to be moved forwardly with respect to the first in the direction of rotation an amount equal to half the circumferential distance between the center of one tooth and the next, and a driven clutch ring having teeth on the face adjacent the driving clutch ring corresponding circumferentially to the teeth of the driving clutch ring but radially as thick as the teeth of the inner and outer driving clutch ring together.

12. The structure defined in claim 11 wherein the driving faces of the jaw teeth are beveled so as to be forced out of engagement by overload and resilient means holding the said jaw teeth in engagement.

13. The structure defined in claim 11 having resilient means urging the faces of the teeth of the driven clutch ring against the faces of the teeth of the driving clutch ring with heavy pressure, and anti-friction means placed to facilitate the said limited rotation of the second driving clutch ring with respect to the first While under said pressure.

14. Power transmission mechanism comprising, a driving member, a driven member, gearing for connecting said members, a jaw clutch for connecting said members directly, a second jaw clutch for connecting said members through said gearing, stressed resilient means for holding said direct drive clutch in engagement,, said direct drive jaw clutch comprising means for forcing itself out of engagement upon overload, and a torque operated means associated with said gear drive jaw clutch for forcing the said direct drive jaw clutch farther out of engagement and holding it so disengaged while gear drive is in efiect.

15. The combination, in power transmission mechanism, of a driving member, a driven member, toothed jaw clutch means on the driving member, corresponding toothed jaw clutch means on the driven member, spring meansnormally pressing the said jaw clutch means into engagement, said jaw clutch means having beveled driving faces whereby sufficient load may overcome said spring means and cause disengagement, and means operative upon said disengagement by the difference in relative rotative speeds of the two members to close the spaces between the teeth of one of the jaw clutch means against entry of the teeth of the other jaw clutch means, whereby said springs maintain a rubbing contact of the ends of the teeth of said jaw clutch means while they rotate at difierent speeds without clatter.

16. The structure defined in claim 15 wherein there is gearing for connecting said members upon said disengagement and torque means associated with said gearing operative after said disengagement and upon assumption of load by said gearing to draw the ends of the said teeth out of the said rubbing contact.

17. The structure defined in claim 15 wherein there is gearing for connecting said members upon said disengagement and torque means associated with said gearing operative after said disengagement and upon resumption of load by said gearing to draw the ends of said teeth out of the said rubbing contact, and centrifugal weight and cam means for increasing the stress in said spring against which said torque means must act at a rate which is less than in direct proportion to the increase in speed of the weights.

18. Power transmission mechanism comprising, a housing, a driving member, a driven member, a ring gear on said driving member, a planet pinion carrier on said driven member, a sun gear concentrically supported within said ring gear,-planet pinions carried on the said carrier in mesh with both the ring gear and the sun gear, a non-rotatable jaw clutch on the housing, a jaw clutch on the sun gear adapted by axial movement of said sun gear to engage the non-rotatable jaw clutch, a jaw clutch secured to the driving member, a second jaw clutch on the driving member adjacent the first but having limited rotation with respect to the first whereby when said rotation is to its limit backwardly of the direction of rotation of the driving member the jaw teeth of the first are midway of the jaw teeth of the second and when said rotation is to its limit forwardly of the direction of rotation of the driving member the jaw teeth of the first and second are in alignment, a direct drive jaw clutch on the carrier having jaw teeth corresponding to the spaces of the driving member jaw clutches and having a face width equal to both of the driving member jaw clutches, springs stressed to normally hold the driving and carrier jaw clutches engaged, said drivingand carrier jaw clutches having beveled driving faces whereby they may be forced out oi. ensseement under load, centrifugal weights rotatable by the carrier, cam means arranged to apply the centrli'ugal force of said weights to said springs, said cams being so formed as to increase the spring stress at a rate which is less than in direct proportion to the speed of the carrier, helical teeth on the sun gear angled to cause/axial movement under load when the direct drive clutch slips to engage the sun gear clutch with the non-rotatable clutch, and a direct drive clutch disengaging member on the carrier operative upon slipping of the direct drive clutch and assumption of the load by the sun gear clutch to be moved by movement of said sun gear into clutch engaging position to completely sever the driving member and carrier clutches.

19. Power transmission mechanism comprising, a driving member, a driven member, gearing for connecting said members, a direct drive jaw clutch for connecting said members directly. a gear drive jaw clutch for connecting said members through said gearing, stressed resilient means holding the teeth of the direct drive Jaw clutch in mesh, said teeth having beveled driving faces whereby they are forced out oi mesh against the force of said resilient means upon overload, means for preventing remeshing of the teeth of the direct drive jaw clutch as long as the driving side revolves i'aster than the driven side, whereby the ends oi the teeth of the driving and driven side -revolve in rubbing contact under pressure of said resilient means, and torqueresponsive means, operative to mesh said gear drive jaw clutch and to overcome said resilient means and withdraw the ends of the teeth of the direct drive jaw clutch out of rubbing contact when said gears assume the load;

20. The structure defined in claim 19 wherein 

